Locomotive power plant



hm, M, 1950 P. s. DICKEY 2,496,041

LOCOMOTIVE POWER PLANT Filed Feb. 15, 1945 2 Sheets-Sheet 1 FIG. 3

II I STEAM a )I l500psi-9OO F |Q AIR INLET FAN ELECTRIC STEAM VAPORGENERATOR TURBINE COMPR L I J VAPOR 6 CONDENSERS 6 STEAM B I coN0 E NsERAMMONIA VAPORIZER 5. FAN

A HOT wlk D WELL E1 LIQUID/ 1-21 8 FIEcE vERI I PRESS. TEMP A LIQUID I5380 I55 B VAPOR I53 80 631 C VAPOR 2I2 I20 649 D LIQUID 2I2 80 I55l'mnentor PAUL 3,. DICKEY Jan. 3], IQEO P, s, DlCKEY ,496fi4flLOCOMOTIVE POWER PLANT Filed Feb. 15, l945 2 Sheets-Sheet 2 PRODUCTSCOMBUSTION FIG. 4

GENERATOR STEAM TURBINE LIQUID RECEIVER OIL COOLER Inventor PAUL S.DICKEY K A/Q l Cittomeg ELECTRIC GENERATOR Patented Jan. 31, 1950LOCOMOTIVE POWER PLANT Paul S. Dickey, East Cleveland, Ohio, assignor toBailey Meter Company, a corporation of Dela- TV are Application February15, 1945, Serial No. 578,009

I 1 Claim. (Cl. 60-95) This invention relates to elastic fluid powerplants and particularly to the apparatus of such plants as used inmobile service, such as for rail way locomotives.

A plant of this type should preferably utilize a steam turbine due toits greater economy and reliability, and it should also use vaporgenerators which will supply the turbine with high pressure and hightemperature steam to obtain the greatest possible economy, from thesteam cycle. The use of the high pressure and high temperature steamrequires thatthe vapor generator be fed with relatively pure and oxygenfree feed water, which means that it is almost essential to use a closedsteam cycle so that feed water is obtained entirely from condensate fromthe steam turbine. It is not practical, however, for such mobile plantsto carry condensing water for the steam condenser. In existing mobilessteam turbine plants an air cooled steam condenser has been used, butthe results with equipment of this type are not entirely satisfactory.

A principal diificulty encountered with aircooled condensers on mobilesteam power plants has been the wide variation in ambient temperature ofthe air encountered by the locomotive and used in condensing the steam.Such a locomotive may be subjected to air temperatures of 110 F. to 40F. Since the air condenser, in order to obtain suficient surface, mustconsist of many small tubes, it is extremely difficult to arrange thecondenser sothat the .water in these small tubes will not freeze .whenthe condenser is subjected to extremely low ambient temperatures. It hasbeen found that the only possible way that freezing can be prevented,particularly if there is any air mixed with the condensing steam insidethe condenser, is to keep a relatively high velocity throughout thesesmall passages. Obviously, such high velocity with the large volumes ofsteam which must be handled requires pressure losses which limits theamount of vacuum available at the turbine outlet.

Furthermore, since the heat transfer per unit of surface is relativelylow many small parallel passages must be used, and these passages beingsubjected to high velocity present a very serious problem ofdistribution of flow. It is apparent that if the flow is not properlydistributed freezing will occur in one section of the condenser, andsince adjacent tubes are subjected to lower air temperatures by virtueof the freezing the tendency is for adjacent tubes to likewise freeze,i. e., for the freezing to migrate across an entire condenser section.

The high velocities which must be used in air cooled condenser passagesin order to prevent freezing make the problem of removing the condensateextremely dimcult, as the tendency is for water condensed to be carriedalong by the steam 2 rather than to drain into suitable receivers forcollecting the condensate. This problem is still further complicated byvirtue of the many sections required for condensing the steam and thediiierent pressures encountered in these sections to obtain the desiredvelocity. Even if suitable collection of condensate could be obtained ineach section of the condenser it is difficult to get this condensate tofiowto one common point properly, since all of these receivers arelikely to be at different pressures, locations and elevations.

The air cooled condenser also makes the problem of air removal quitedifiicult, since there is no common point where condensate may becollected and deaerated as is the case in the conventional design ofliquid cooled steam condensers.

Furthermore, the air cooled surfaces must be located along the exteriorwalls of the locomotive, usually not adjacent the turbine, thusnecessitating a complicated and obstructing system of the large pipesnecessary to conduct the exhaust steam to the condenser and of thesmaller conclensate returns.

If freezing temperatures, or wide variations in ambient temperature, arenot to be encountered then the field of travel of the mobile unit isgreatly restricted. For example, trans-continental runs, encounteringdesert temperatures and mountain snows, must be avoided.

At the other temperature extreme, namely, continuous operationencountering high ambient temperatures, excessively large heat transfersurfaces are required for air cooling a steam condenser, with consequentincrease in piping and fan power.

It is a principal object of my invention to provide improvements inapparatus and method of operationwhereby the disadvantages of known aircooled steam condensers are obviated and a higher heat cycle elficiencyis obtained for any given vapor generator-vapor turbine installation.

A particular object is to provide an improved cooling system forcondensing the steam ex lgausted from aturbine in mobile service wherethe usual arrangements of condenser cooling water are not available.

In the drawings: Fig. 1 is a diagrammatic showing of one embodiment ofmy invention.

wFig. 2 is a tabulation of fluid conditions in connection with Fig. 1.

Fig. 3 is a diagrammatic elevation of a portion of a locomotive.

Fig. 4 is a diagrammatic representation of anotherembodimeht' ofrrii,invention.

The'conomy of a mobile steam turbine een operating at 1500 p. s.hand-900 F total tem perature and having air cooled condensers,operating under most favorable conditions of a clean condenser andrelatively low to F.) air temperatures, may be little better than thatof a non-condensing steam engine cycle. If suitable vacuum could bemaintained at the turbine exhaust the horsepower output of the plantwould be materially increased and at a water rate approaching beststationary practice.

I have mentioned that in an attempt to avoid freezing of an air cooledcondenser the economy of the total plant suffers through pressure dropvelocities and other attendant factors. One known way of eliminating thefreezing problem is to use a bifluid condenser which would consist of asteam condenser of the conventional type connected to the turbineexhaust and using a coolant of the non-freezing type, such as Prestoneor any or" the other common coolant-s used in gasoline engines. Thecoolant would then flow to the air cooled heat exchanger where the heattaken up by the coolant in the steam condenser is removed by the air.

Such a system has the obvious advantage that difficulties due tofreezing are eliminated, since the steam condenser can be placed in aprotected zone underneath the turbine and the coolant, which isnon-freezing, is the only medium subjected to low ambient airtemperatures. It has the further advantage that the long largedlameter(because of low pressure and large volume) lines which must be used inthe case of the air cooled steam condenser are eliminated.

However, this known system has several disadvantages which makes thescheme rather impractical. Since the system must work on a lowtemperature dii'ference suitable vacuum is to be maintained at theturbine exhaust, a very quantity of coolant is required. On ordinarysteam condensers the quantity of cooling water at relatively lowtemperature 60 F.) for vacuums of the order of 28 inches of mercuryaverage from sixty to one hundred times by it the amount of steamcondensed. This me; that under the most favorable conditions a plantusing 50,000 lb. of steam per hour would have to recirculate about5,000,000 lb. of coolant per hour. Furthermore, the temperaturedifference must be used twice. That is, if the plant is designed to usea 20 temperature difference in the heat exchangers and an airtemperature of 100 F. were encountered, a minimum coolant temperatureentering the steam condenser would be 120 F. and the minimum condensatetemperature would be 140 F., which corresponds to a vacuum of only 23inches in place of the 28 inches desired.

A further disadvantage is that the heat transfer rate in the air cooledheat exchanger is low, so that extremely large surfaces would berequired in fact considerably larger than would be required for thedirect air cooled steam condenser.

The bifluid condenser system however does have the advantage ofelimination of the freezing problem and elimination of the problem ofhanling condensate and entrained air, and also the elimination ofextremely large pipe lines between the turbine and air condenser, whichmust of necessity be located some distance away in order to arrange thesurface so that the cooling air can be passed through it. Against theseadvantages are the distinct disadvantages of the tremendous quantity ofcoolant which must be pump circulated, the size of the air cooled heatexchanger, and the loss-of vacuum when higher ambient air temperaturesare encountered.

My present invention provides a combination of apparatus and method ofoperation overcoming the mentioned disadvantages of the air cooled steamcondenser or coolant steam condenser. It is in simplest terms theapplication of a refrigoration cycle to the condensing system of a highpressure steam power plant. Fig. 1 shows diagrammatically the principalapparatus of such a system as well as the flow piping for both the steamcycle and the refrigerant cycle. Fig. 2 is a tabulation of refrigerantvalues which may exist at locations indicated on Fig. 1.

In the particular example which I have taken for consideration I utilizean electrically propelled locomotive having an electric generator Idriven by a steam turbine 2. Approximately 50,000 lb./hr. of steam at1500 p. s. i. and 900 F. total temperature is supplied the steamturbine. If the turbine were operating at a back pressure of 28 inchesof vacuum, such a plant should produce approximately 6500 H. P. at awater rate of '7 .5 lb. per horsepower. I will show that through the useof my invention the system will approach this output, and even afterdiscounting the increased power required for auxiliaries there remainsan increase in output horsepower of 60 92 to 70% over that which may beexpected with an air cooled steam condenser as previously discussedherein.

The turbine 2 exhausts to a steam condenser 3 having a hot well 4 fromwhich the condensate is returned as feed water to the vapor generatorsupplying the turbine. In this arrangement the steam condenser islocated immediately adjacent the steam turbine, thus eliminating theconsiderable runs of very large piping previously necessary to carry theexhaust steam to conventional side wall air cooled condensers usuallylocated on opposite sides of the locomotive, as indicated at 5 in Fig.3. Such air cooled condensers 5 are normally provided with a centrallylocated exhaust fan 6 discharging through the roof of the locomotivethrough louvres l. The power plant of this invention is particularlyadapted for installation in a locomotive of the type illustrated in Fig.3.

For heat transfer in the steam condenser I utilize a refrigerant,preferably ammonia. While it is apparent that any of the commonrefrigeration cycles, such as dense air compression, vapor compression,or continuous absorption types may be used, the vapor compression systemwhich is used extensively for refrigeration work at the present timeseems to best apply to this problem, and I have chosen as an example theuse of ammonia. Thus in my preferred embodiment of Fig. l the steamcondenser 3 becomes an ammonia vaporizer.

While there is a wide range of refrigerants available for the range oftemperatures which would be used in this application, many of theserefrigerants have obvious disadvantages, such as fire hazard orcorrosive action. However, it appears that ammonia, methyl ohloride andFreon all satisfy the conditions of the cycle desired and do not presentany particularly severe handling problem. Only very slight modificationsof an ordinary steam condenser will be necessary in order to preventpollution of the condensate with the refrigerant in case leaky tubesshould develop, but this does not present any serious difficulty. It isentirely possible that a refrigerant operating at low pressure could beused if desired, but it appears that the most economical refrigerant isone which operates at somewhat higher pressure, and thus lower specificvolume.

The media generally used in compression machines, ammonia, sulphurdioxide, carbon dioxide, etc., exist only as a gas or vapor atatmospheric pressures and ordinary temperatures, but they are liquefiedWhen compressed to a sufiiciently high pressure and cooled. The heatabsorbed in reevaporating the liquid at a reduced pressure constitutesthe refrigerating effect. To periodically return the refrigeratingmedium to its original liquid state the system must include thefollowing parts:

1. Evaporating space wherein the liquid is evaporated, absorbing heatfrom its surroundings and producing the refrigerating effect.

2. A compressor in which vapor from the evaporating space is compressedand supplied the 'condenser at a terminal pressure corresponding to thetemperature of the saturated vapor obtainable with the cooling effectavailable.

3. The condenser in which the latent heat and heat of compression isremoved and the vapor is liquefied by air passing over the condensertubes.

ammonia from a liquid receiver 8 passes through an adjustable expansionvalve 9 to the tubes of the steam condenser 3 in which the ammonialiquid is vaporized.

The ammonia vapor leaving the steam condenser 3 passes to a vaporcompressor l0 driven by the steam turbine. In the compressor the ammoniavapor has its pressure increased and gains a slight amount of heat ofcompression. The vapor at higher pressure passes to one or more heatexchangers H which are air-cooled and are, in fact, ammonia vaporcondensers. Here the ammonia vapor at substantially the same elevatedpressure loses its heat of vaporization plus the superheat it gained inthe compressor and passes as a liquid to the liquid receiver. Thetabulation Fig. 2 gives general values of pressure, temperature and heatcontent for the liquid or vapor ammonia which may be expected atlocations A, B, C and D in the refrigerating cycle.

The system has, among others, the following advantages over theair-cooled steam condenser or the bifiuid condenser systems previouslymentioned:

l. The problem of freezing in the air cooled stream condenser isdefinitely eliminated, as the steam condenser is here placed in aprotected location adjacent the turbine and the fluidpassing through theair-cooled heat exchanger will not freeze under any ambient temperatureconditions encountered.

2. The pressure loss which is encountered in the exhaust system of theair cooled steam condenser is eliminated so that good vacuum isavailable at th biee 3. A conventional design of steam condenser may beused which eliminates the problems of air removal and condensatedrainage.

4. The design of the air cooled heat exchanger is very materiallysimplified as the problem of cooling the refrigerant is not complicatedwith the problems of maintaining vacuum, removal of noncondensibles, andsimilar problems inherent in the steam condenser. A possibility ispresent of dividing the air cooled heat exchanger into a numberofsections and placing these atmost convenient locations around thepower plant. The piping for the vapor or liquid refrigerant to and fromthe air cooled heat exchangers is very decidedly smaller than to conductthe exhauststeam to an air cooled steam condenser.

5. This cycle takes advantage of the high coefiicient of heat transferfrom a boiling or con.- densing fluid and has the advantage of providinga satisfactory temperature difference in both heat exchangers withoutloss of vacuum. at..the turbine exhaust. i 1 1.

6. A relatively small quantity of refrigerant must be recirculated sincethe latent heat of vaporization is utilized. Instead of. requiring morethan one hundred pounds of coolant per pound of steam condensed, as is.the=case. of .the bifluid condenser, the system will require only twoto ten pounds of refrigerant per pound' of steam condensed.

7. The smaller quantity of refrigerant required and the reduced volumeof the refrigerantrequires very much smaller piping to-andfrom the heatexchangers. For example. in the .air cooled steam condenser with theturbine operating at 28 inches of vacuum,-- roughly 400,000 C. F. M. ofair must be'passed tothe air cooled heat exchanger. In the bifluidcondenser ap proximately 5,000,000 lbs. per hour of liquid coolant mustbe recirculated throughvthe air cooled heat exchanger. -In the proposedrefrigeration cycle only about 3600 C. F: of ammonia vapor must becirculated through the heat exchanger:

8. Stoppage ihany portion of the air cooled heat exchanger does notseriously affect the performance of the system.

9. The oil coolers for the turbine-generator lubricating oil may bebuilt into the refrigeration system, which is an advantage inasmuch asthe latter would normally be located near the steam condenser and theoil and refrigerant piping would be quite simple.

10. Control of the refrigerating system to obtain the maximum possiblevacuum with a minimum amount of compressor horsepower could be veryeasily accomplished with a thermally operated valve in the refrigerantline.

It will be appreciated that the various values which I have used in thisdescription are by way of example only. I have not attempted to give acomplete heat balance for any of the. cycles or systems discussed, asthis appears to be unnecessary and would obviously represent only aparticular set of conditions in any event. I have arbitrarily chosen tobase my discussion upon a locomotive having a vapor generator capacityof approximately 50,000 lb./hr. of steam at 1500 p. s. i. and 900 F.total temperature. Ordinarily this should produce a turbine output ofapproximately 6500 H. P. with a steam condenser operating at '28 inchesof vacuum. Allowing even as much as 500 H. P.for the exhaust fans of theair cooledheatexchangers and as high as 1000 for the aeemr ssea: meangain in horsepower is 50-70% over a non-condensing steam enginelocomotive or an air-cooled steam condenser turbine installation.

While in Fig. l I have diagrammatically indicated the vapor compressoras directly driven from the steam turbine, it will be appreciated. thatthis may be through the necessary speed reduction gears, but thelocation of the compressor will obviously be'adjacent the steam turbineand steam condenser to take advantage of the shortest possible runs ofpiping for the ammonia vapor. Also the compressor being driven by themain turbine increases the size of the main turbine, thus making it moreefficient.

This preferred cycle takes advantage of the high coeihcient of heattransfer from the b-iling or condensing fluid, that is the heat transferbetween condensing steam and boiling ammonia.

In Fig. 4 I follow the same general arrange ment as in Fig. 1, exceptthat I have indicated that a portion of the air cooled heat exchangersurface may be located in the path of the forced draft supply for thevapor generator. The vapor generator is indicated at [2 and includes aconventional combustion chamber. It is supplied with any convenientfuel, such as oil or pu1- verized coal, and with air for combustionproduced by a fan is passing air under pressure through a conduit M. Inthe conduit i4 is located a secondary heat exchanger l5 through which aportion of the high pressure ammonia vapor is passed for condensing thesame. The heat taken from the ammonia vapor, namely, the latent heat ofvaporization plus superheat due to the compressor is given oif to thecombustion air, thus aiding the efficiency of the combustion processwithin the vapor generator H2. The heat exchangers i l comprise primaryheat exchange or condensing means.

For cooling the bearing lubricating oil of the steam turbine I indicatethat the oil may be passed through a coil I5 located in the liquidreceiver 8. At this location the ammonia liquid (for the presentexample) is under pressure of approximately 212 p. s. i., a temperatureof 80 F. and a B. t. u. content of 155 B. t. u. per pound. The oiltemperaure being in the neighborhood of 140 F. and of only a fractionvolume flow rate compared with the volume flow rate of the liquidammonia will be cooled toward 80 F. with no very appreciable rise intemperature or B. t. u. content of the ammonia liquid. Here again thelocation is ideal, as the liquid receiver is preferably located withsubstantially nothing other than the expansion valve 9A between theliquid receiver and the inlet to the ammonia vaporizer 3. This locationin immediate proximity to the steam condenser 3 and steam turbine 2 minimizes the piping for turbine bearing oil to and from the oil coolingcoil I6.

Control of the system to obtain maximum possible vacuum with the minimumamount of compressor horsepower may readily be accomplished with athermally operated expansion valve 9A in the liquid ammonia pipe leadingfrom the liquid receiver 8 to the ammonia vaporizer 3. A temperaturesensitive bulb ll automatically actuates the expansion valve 9A. Ifdesired, this may be a pressure responsive expansion valve sensitive topressure within the hot well 4, which pressure is, of course, a functionof temperature under these absolute pressure values of high vacuum.

In applying the refrigerating cycle to the steam condenser of such apower plant there is the further possibility of making this cycle workas a binary vapor engine, as well as producing the necessar cooling forthe steam condenser. If the compressor were designed so that it wouldfunction either as a compressor driven from the main turbine, or as aturbine putting power back into the main unit, the system itself couldbe arranged so that it would either add to or subtract from the power ofthe main turbine. It might be desirable to have an ammonia vapor turbineand an ammonia vapor compressor as separate units, with the necessaryvalving.

If the ambient air temperature were high it would then be necessary tofurnish power from the main turbine to the compressor in order toproduce the necessary refrigerating action to give the desiredtemperature difierence both in t e steam condenser and in the air cooledheat exchanger.

If, however, the ambient temperature were low enough, then thereirigerating act.0n would not be necessary and a lower pressure couldbe maintained in the air cooled heat exchanger than in the steamcondenser, and the action would be for the vapor generator in the steamcondenser to run the compressor and furnish power back to the main unit,this vapor being condensed in the air cooled heat exchanger. A liquidpump would obviously be necessary in this case, and this pump could beconnected in the liquid line and run only at times when the ambienttemperature is low enough to use the compressor as a power generator.The horsepower required in the liquid pump would be quite small, sincethe quantities circulated are not large in liquid phase.

While I have chosen to illustrate and describe only a preferredembodiment, it will be understood that this is by way of example only.The type of refrigerant suggested, as well as the various vaiues ofquantities, temperatures, pressures, etc, are given by Way of examplerather than as indicative of a definite heat balance.

What I claim as new, and desire to secure by Letters Patent of theUnited States, is:

A turbo-electric railway locomotive comprising a steam generator havinga combustion chamber, a steam turbine connected with said steamgenerator, a steam condenser arranged immediately adjacent said turbineand connected to receive exhaust steam therefrom, said condenser beingarranged within the locomotive for protection from the exteriortemperature conditions, a closed cycle volatile liquid refrigerantsystem comprising primary refrigerant condensing means subjected totemperature conditions at the exterior of the locomotive, secondarrefrigerant condensing means disposed in the path of the draft supply tothe combustion chamber for the steam generator, passage means connectingboth said refrigerant condensing means in parallel with said steamcondenser, a vapor compressor connected in said passage means at theintake side of said refrigerant condensing means, a liquid receiverconnected in said passage means at the discharge side of saidrefrigerant condensing means, and an expansion valve arranged in saidpassage means between said liquid receiver and said steam condenser.

PAUL S. DICKEY.

(References on following page) REFERENCES CITED The following referencesare of record in the file of this patent:

UNITED STATES PATENTS Number Name Date Dodge Jan. 20, 1891 Murray Aug.30, 1898 Patten July 9, 1912 Watson Oct. 30, 1934 Fisher Mar. 14, 1939Larrecq June 18, 1940 Number Sarco Temperature Regulator, in

10 Name Date Schwarz Nov. 4, 1941 Price Aug. 25, 1942 FOREIGN PATENTSCountry Date Switzerland Apr. 1, 1924 OTHER REFERENCES The Chemical Age,issue of Nov. 4, 1944, pages 447 and 448.

